Pump apparatus

ABSTRACT

In first and second forms of variable-delivery reciprocating pumps wherein fluid delivery past a first spring-loaded delivery check valve is provided, delivery per stroke is decreased with increasing pump drive speed by provision of a second springloaded check valve having a smaller inertia-to-spring force ratio than that of the delivery check valve, so that the second check valve is temporarily opened by peak pressures occurring in the cylinder to release fluid from the pump cylinder. Delivery is basically controlled in the first pump by controlling the point during the stroke when the piston closes off an inlet port, and in the second pump by controlling the point when the cylinder is connected to a spill-back conduit, and in either pump the basic control also may vary the spring loading on the second check valve. Variation of the spring loading on the second check valve allows one to render the characteristic of the first pump more or less dependent upon the delivery control setting. In a preferred third variable-delivery reciprocating pump, delivery is initiated by closing an inlet port and opening an outlet port with an overlap during which a finite minimum port area for fluid flow from the pump cylinder causes a finite maximum restriction to flow from the cylinder and a peak pressure within the cylinder. A spring-loaded check valve responsive to cylinder pressure diverts fluid from the cylinder. Delivery is basically controlled by adjusting the points during the stroke at which the inlet port is closed and the outlet port opened. By suitably arranging the port geometry, the finite minimum port area occurring during the stroke may be made to automatically vary with the basic delivery adjustment in accordance with a desired relationship.

United States Patent 1 1 Ulbing [451 Oct. 1, 1974 1 PUMP APPARATUS OtmarM. Ulbing, Owego, NY.

[73] Assignee: Borg-Warner Corporation, Chicago,

[22] Filed: Mar. 1, 1973 [21] Appl. No.: 336,983

Related US. Application Data [60] Division of Ser. No. 155,114, June 21,1971, Pat. No. 3,739,809, which is a division of Set. No. 857,162, Sept.11, 1969, Pat. No. 3,614,944, and a continuation-impart of Ser. Nos.786,233, Dec. 23, 1968, abandoned, and Ser. No. 153,460, June 1, 1971,Pat. No. 3,740,172.

[75] Inventor:

Primary ExaminerWilliam L. Freeh Attorney, Agent, or Firm-Richard G.Stephens 57 ABSTRACT In first and second forms of variable-deliveryreciprocating pumps wherein fluid delivery past a first spring- SPLYloaded delivery check valve is provided, delivery per stroke isdecreased with increasing pump drive speed by provision of a secondspring-loaded check valve having a smaller inertia-to-spring force ratiothan that of the delivery check valve, so that the second check valve istemporarily opened by peak pressures occurring in the cylinder torelease fluid from the pump cylinder. Delivery is basically controlledin the first pump by controlling the point duringtlhe stroke when thepiston closes off an inlet port, and'in the second pump by controllingthe point when the cylinder is connected to a spill-back conduit, and ineither pump the basic control also may vary the spring loading on thesecond check valve. Variation of the spring loading on the second checkvalve allows one to render the characteristic of the first pump more: orless dependent upon the delivery control setting. In a preferred thirdvariable-delivery reciprocating pump, delivery is initiated by closingan inlet port and opening an outlet port with an overlap during which afinite minimum port area for fluid flow from the pump cylinder causes afinite maximum restriction 'to flow from the cylinder and a peakpressure within the cylinder. A springloaded check valve responsive tocylinder pressure diverts fluid from the cylinder. Delivery is basicallycontrolled by adjusting the points during the stroke at which the inletport is closed and the outlet port opened. By suitably arranging theport geometry, the finite minimum port area occurring during the strokemay be made to automatically vary with the basic delivery adjustment inaccordance with a desired relationship.

OP TO NOZZLE mmmum mm 3I88894 I MINIMUM DELIVERY PUMP APPARATUS Thisapplication is a division of my prior copending Application Ser. No.155,114 filed June 21, 1971 now U.S. Pat. No. 3,739,809, which in turnis a division of my prior Application Ser. No. 857,162 filed Sept. 11,1969, now U.S. Pat. No. 3,614,944, issued Oct. 26, 1971, and acontinuation-in-part of my prior applications Ser. No. 786,233 filedDec. 23, 1968, now abandoned, and Ser. No. 153,460 filed June 1, 1971now Pat. No. 3,740,172.

It is desirable, after starting, and during normal warmed up operations,that the fuel'to-air ratio of the mixture burned by an engine eitherremain relatively constant or vary slightly in a desired sense over awide range of speed and load conditions. It is generally preferred thatthe ratio either remain constant or vary slightly inversely with enginespeed. Most internal combustion engines are controlled by a primarycontrol which operates basically to control engine torque. In mostcarburetor-equipped engines the engine primary control controls torqueby adjusting a throttle plate to adjust the amount of fuel-air mixtureaspirated into an engine cylinder during a stroke. The amount of mixtureaspirated per stroke determines the amount of air drawn into thecarburetor air intake per stroke, and because of the venturi principleupon which the carburetor operates, the amount of fuel mixed with airvaries rather proportionally. Because carburetor air flow inherentlycontrols carburetor fuel flow, the fuel-to-air ratio of the mixtureinherently tends to remain somewhat constant over a wide range of enginespeeds. Any unwanted variation can be corrected easily in a number ofways, such as by use of a metering needle valve moved in response toengine vacuum.

While fuel injection engine systems have a number of advantages overcarburetor-equipped engine systems for a number of applications, fuelinjection systems do not have the above-described automaticmixturemaintaining tendency of carburetor systems. In fuel injectionsystems engine torque is usually controlled by simultaneous variation ofthe amount of fuel pumped per stroke and the amount of air aspirated perstroke, with the engine torque control linked to both vary the amount offuel pumped per stroke and to vary a throttle plate in the air intakeduct. When a decrease in engine load causes an increase in engine speedat a given torque control setting, the increased air flow causes anincreased pressure drop across the air intake structure, decreasing theamount of air aspirated per stroke. With the amount of air aspirated perstroke decreased and the amount of fuel pumped per stroke remaining thesame, it will be seen that the fuel-to-air ratio disadvantageously tendsto increase as engine speed increases due to decreased load at a giventorque control setting. While such a variation to fuel-air ratio is notregarded as a serious limitation in many applications, in certain otherapplications, and particularly in those where the engine frequentlyoperates under widely-varying load conditions, it is desirable that fueldelivery vary inversely with engine speed, or directly with engine load,as well as with adjustment of the engine torque control, or primarycontrol. The variation of fuel delivery so as to maintain a desiredfuel-to-air relationship may be termed secondary control. It is aprincipal object of the invention to provide improved pump apparatushaving effective and reliable means which provide pump operatingcharacteristics such as those useful for automatic secondary control.

Secondary control, or automatic variation of fuel delivery with enginespeed in order to keep the fuel-air mixture relatively constant, isknown in the prior art in connection with diesel engines and certaingasoline engines which use fuel injection. One prior art system is shownin U.S. Pat. No. 3,443,554, for example. However, those prior artsystems of which I am aware require a centrifugal governor and/or othervery complex and expensive mechanisms in order to provide such pumpdelivery versus speed characteristics. A very important object of theinvention is to provide such pump characteristics in a reciprocatingpump using much simpler and more economical apparatus which is easilyconstructed and highly reliable. Use of the present invention allows oneto replace extremely complex and expensive mechanisms with merely aspring-loaded check valve.

In some prior art pumps used in fuel injection systems, the maximumpressures developed in the injection pump during an engine cycle varyappreciably with both engine speed and the fuel delivery setting of thebasic torque control, while the maximum pressure developed in variousother systems varies substantially solely with engine speed and issubstantially independent of the fuel delivery setting of the basictorque control. As will be seen below, desired automatic control of somereciprocating pump systems may be obtained solely by use of aspring-loaded check valve having a constant spring load, and desiredautomatic control of the other prior systems may be obtained very simplyby use of a spring-loaded check valve together with means for varyingthe spring load as a function of the setting of the basic torquecontrol. Thus further objects of the invention are to provide improvedreciprocating pumps of both types having desired delivery versus speedcharacteristics.

Several basically-different types of variable-delivery reciprocatingpumps are known, and used in fuel injection systems, for example, andanother object of the invention is to provide delivery versus pump speedcontrol arrangements which are useful with each of the different typesof pumps.

A principal object of the present invention is to pro vide an automaticdesired variation of pump delivery with pump drive speed.

Other objects of the invention will in part be obvious and will in partappear hereinafter.

The invention accordingly comprises the features of construction,combination of elements, and arrange ment of parts, which will beexemplified in the constructions hereinafter set forth, and the scope ofthe invention will be indicated in the claims.

For a fuller understanding of the nature and objects of the inventionreference should be had to the following detailed description taken inconnection with the accompanying drawings, in which:

FIG. 1 is a cross-section view of a reciprocating variable-delivery pumpof the type shown in my prior application Ser. No. 786,233 modified toincorporate a secondary control delivery versus speed characteristic inaccordance with the present invention.

FIG. la shows a modification to the pump of FIG. 1 to provide adifferent type of oil pumping.

FIG. llb is a porting diagram useful in understanding the operation ofthe pump of FIG. 1.

FIGS. 2a, 2b and 2c are diagrammatic views useful in explaining theoperation of the invention with each of three different types ofreciprocating. variabledelivery, constant stroke-length fuel injectionpumps.

FIG. 2a diagrammatically illustrates the application of the invention toa well-known form of fuel injection system using a variable-deliveryreciprocating pump in which delivery is varied by varying the time orpoint during the stroke at which an inlet port is closed off. The enginecrankshaft is mechanically connected to reciprocate piston P a fixedstroke distances within cylinder CY. Piston P is provided with an axiallength which varies angularly around the piston, so that rotating pistonP to various angular positions will vary the time or position duringeach stroke at which the piston covers inlet port lP. lf port IP isclosed off early during the stroke, greater delivery will result. Duringa rightward pumping stroke, fuel will be expelled through inlet port IPback to the supply tank until the piston blocks the port, and forwardpumping will occur through delivery check valve DV to nozzle NO afterthe piston blocks the port until the end of the stroke. The angularposition of piston P is controlled by the engine primary control, whichis shown as comprising an accelerator pedal A. The primary control alsovaries air intake throttle plate TP via a cam or suitable linkage L togenerally increase air flow as fuel quantity is increased. At a givensetting of the 'primary control, an increase in engine speed due to adecrease in load will decrease the amount of air aspirated per strokedue to the restriction of the air intake structure S, therebyundesirably increasing the fuel-air mixture ratio.

As piston P of the injection pump travels rightwardly on a pumpingstroke from the leftward limit position shown in FIG. 2a, fluid willinitially be expelled out inlet port [P If piston velocity were uniform,the pressure within the cylinder during the initial travel would tend tobe a constant value dependent upon piston speed and the restriction ofunblocked port IP. If piston P is instead reciprocated with simpleharmonic motion, or an approximation or modification thereof, as isusually the case in practical systems using cranks or eccentrics or thelike, the linear velocity of the piston will instead vary approximatelycosinusoidally, for example, from zero velocity at the leftward positionshown to a maximum speed at mid-stroke, down to zero velocity at the endof the rightward stroke. The linear velocity, assuming simple harmonicmotion, may be written as (taking mid-stroke as the origin):

hence the pressure will increase approximately in accordance with anaflcoswt) characteristic, where w is pump drive speed in radians persecond. Thus the pressure existing in the cylinder when the piston edgereaches the inlet port will be higher at higher pump drive speeds,varying approximately with the'square of pump drive speed. If the area Aof port IP is large,

the pressure built up in the cylinder prior to closure of the inlet portwill still be very modest, however, and in many systems is small enoughto be neglected. P If at a given phiii Ziriiibed "(HE E8651; rotated."to close off port lP later during the stroke, so as to pro-; videlesser delivery per stroke, but still prior to mid-; stroke, the pistonwill be seen to reach a greater linear; velocity by the time it beginsto close off port IP. and hence a greater pressure will exist whenclosure beginsi than at greater delivery settings. At any given pumpdrive speed, the maximum pressure would be develloped if port [P isclosed off approximately at midstroke. when the piston is at its'maximumlinear velocity. lf piston P is rotated to provide port closure verylate during the stroke, a lesser peak pressure will be developed due .tothe lesser piston velocity at the time of closure. Prior art systems ofthe type shown in FIG. 2a did not include the further check valve SV.and its presence should be ignored for the moment.

As the piston begins to close off port lP, the pressure furtherincreases, not only due to the increase in piston linear velocity andincrease in flow, (assuming port closure prior to midstroke) but alsodue to the decrease in unblocked port area irrespective of whether portclosure occurs before or after midstroke. The pressure p rises in thecylinder as the port is closed off roughly in accordance with thefollowing characteristic:

is thereafter limited. The rate at which the pressure rises as the portis closed off depends not only upon the geometry of the port, but alsoupon the piston speed. The increase in piston speed not only increasesthe pressure due to provision of increased flow, but also increases therate of pressure increase by more quickly closing off the port, so thatthe rate of pressure increase varies as a fairly high power of pumpdrive speed. The

precise slope of the pressure characteristic will also de-' pend, ofcourse, upon the shape of the inlet port as well as its general width,and the shape of the piston edge which passes over the port to block theport. As the piston is rotated to decrease the delivery per stroke at agiven pump drive speed, so that the piston has a greater velocity whenit closes off the port, the rate of pressure increase will be seen toincrease. Since maximum linear piston velocity occurs at midstroke, themaximum peak pressure for any given pump drive speed will occur when theport is closed off approximately at midstroke. when the pump is adjustedto pump approximately onehalf its maximum delivery per stroke.

The rapid increase in pressure as the port is closed off applies asudden force to the body of valve DV, accelerating it rightwardlyagainst the force of the valve spring and providing a damped oscillationof the valve body. The mass of the valve body, the valve spring, and

the viscous resistance of the fuel to motion of the valve body after thevalve is opened will be seen to provide a mass-spring-damper secondorder system. The valve body, eventually returns to a steady-stateposition such that cylinder pressure balances the valve spring loading.and cylinder pressure remains substantially at the value determined bythe delivery valve spring loading for the remainder of the rightwardpumping stroke. The motion ofthe check valve body required to allowmaximum flow through the check valve is assumed to be small compared tothe length of the check valve spring, and hence the spring may beassumed to apply a substantially constant force to the check valve body.Because the pressure drops quickly when the delivery check valve opens,the force applied to the valve body has the nature of a brief impulse,the amplitude of which varies as an exponential function of pump drivespeed. Thus increased pump drive speed increases the amount which thevalve body overshoots. After the overshoot, the check valve maintainscylinder pressure substantially at a value determined by the check valvespring loading. The pressure in the cylinder may increase slightly up tomidstroke as piston velocity increases and thereafter decrease somewhatas piston velocity decreases during the latter half of the rightwardpumping stroke but no further sudden increase in pressure will occurduring the pumping stroke. As men tioned above, the foregoingdescription of operation assumes that check valve SV is not present.

In accordance with the present invention, pump delivery per stroke maybe decreased with increasing pump drive speed from its otherwise valueby provision of the further secondary control check valve SV, which isresponsive to pump cylinder pressure and operative to spill backincreasing amounts of fuel to the supply as pump drive speed increases.The ratio between the mass of the body of a check valve to the springforce of the spring of the check valve may be termed the check valvetime constant. Check valve SV is provided with a smaller time constantthan that of delivery check valve DV. As the closure of inlet port IPcauses the rapid increase in cylinder pressure, the pressure is appliedsimultaneously to both the delivery check valve and check valve SV. Thepressure temporarily rises above the steady-state delivery valvepressure setting due to the greater inertia or longer time constant ofthe delivery check valve, which delays its opening. During thattemporary high pressure condition the secondary control check valve SVopens, due to its lesser mass, despite its greater spring loading, andopening of valve SV spills back fluid to the supply tank and limits thepressure developed in the cylinder. The amount which valve SV opens willbe seen also to depend upon the peak pressure developed in the cylinder,and hence upon pump drive speed. As well as spilling back some fluid,the quick opening of pressure-sensitive valve SV also decreases the timewidth of the pressure impulse applied to delivery check valve DV,thereby decreasing the overshoot of valve DV.

The variation in volumetric efficiency, or air aspirated per stroke,with speed is ordinarily non-linear for most engines. Also, thevariation of peak pump cylinder pressure with pump speed is nonlinear,and the variation in the amount of fuel which a typical springloadedcheck valve will pass with a given pressure impulse applied to it isalso non-linear. Furthermore, the peak pump cylinder pressure occurringat a given engine speed or pump drive speed varies in accordance withfuel delivery setting, as described above. Because of these varyingrelationships, it is sometimes difficult to provide a desired fuel-airratio characteristic over widely-varying load conditions ifa fixedspring loading is used on the check valve SV. lln accordance with afurther feature of the invention. the loading on the secondary controlcheck valve SV may be varied as a function of the primary controldelivery setting, and in FIG. 2a cam C rotated by the primary control Ais effective to vary the spring load on check valve SV.

If the piston in FIG. 2a is rotated to decrease fuel delivery by closingoff the inlet port. later during the first halfof the stroke. the pistonwill have a greater velocity as it closes off the port. therebyincreasing the slope of the pressure characteristic, as will be apparentfrom expression (2), and thereby providing greater impulses to open thesecondary check valve SV. Piston velocity decreases during the latterhalf of the pumping stroke. Thus maximum peak pressure for a givenengine speed is developed if the inlet port is closed approximately atmidstroke, which results when the pump is operating at roughly one-halfof its capacity. Most engine systems require fuel delivery which variesfrom none or some small amount up to a maximum required for normalrunning, although even greater delivery may be required for starting.Since minimum delivery requires inlet port closure very late in thestroke, a pump of the type shown in FIG. 2a when used in such an enginesystem ordinarily will operate over a range which varies from a minimumdelivery condition involving port closure very near the end of therightward stroke when piston velocity is low, up to a maximum deliverycondition involving port closure much earlier during the stroke whenpiston velocity is greater. If themaximum fuel required by the engineduring running conditions is no more than half the maximum pumpcapacity, it will be seen that the peak pump cylinder pressure developedat a given engine and pump speed will vary directly. though notlinearly, with the fuel delivery setting over the entire running rangeof the engine. Under such con ditions, cam C will ordinarily provide aspring-loading to valve SV which generally increases as the delivery isincreased. If the engine requires more fuel delivery than half the pumpcapacity so that inlet port closure must occur prior to midstroke, aplot of the peak pump pressure developed at a given engine speed versuspump delivery will be seen to slope downwardly at the highest deliveryvalues. Under such an arrangement. cam C will ordinarily providespring-loading which increases as delivery is increased up to a givendelivery value, after which cam C will provide decreasing springloadingas delivery is further increased.

FIG. 2b diagrammatically illustrates a different form of reciprocatingvariable-delivery constant strokelength injection pump in which deliveryis varied by varying the time during the stroke at which forward pumpingis terminated, rather than varying the time at which it begins. Thepistons P and AP are reciprocated by the engine with some approximationof simple harmonic motion. As piston P travels rightwardly on a pumpingstroke, delivery commences substantially immediately through deliverycheck valve DV, and continues throughout the rightward pumping strokeuntil port TP of auxiliary piston AP registers with port SP of collarCO, at which time fuel is spilled back through hose H to the supplytank. The delivery check valve feeds a nozzle extending into the engineair intake structure in the same manner as in FIG. 2a. The piston Pcontains a bore and a conduit communicating with port TP of auxiliarypiston AP. Collar CO is arranged to be axially adjustable relative toauxiliary piston AP by means of the engine primary control. so that thetime or position during the stroke at which forward pumping ceases maybe varied to vary the quantity of fuel delivered. Inlet check valve IVadmits fuel to the pump cylinder during the leftward return or suctionstroke.

First consider the operation without the use of secondary control checkvalve SV. At the beginning of a rightward pumping stroke, piston speedbegins at zero and increases cosinusoidally. Pressure builds up in thepump cylinder substantially immediately to a value greater than thesteady-state spring loading of the delivery check valve, and thendecreases to a value commensurate with the delivery check valve loading,as the delivery check valve DV opens. While the velocity of the piston Pis minimum (zero) at the beginning of the stroke, the acceleration ofthe piston is then at its maximum value, and assuming simple harmonicmotion: a s/2 sin cut. The maximum acceleration of piston P applies amaximum impulse to delivery check valve DV, and the magnitude of theimpulse will be seen to vary as the square of engine and pump speed. Asthe delivery check valve opens, the pressure in the cylinder dropsmarkedly. The pressure then increases somewhat until midstroke (assumingcollar CO is adjusted to provide delivery past midstroke) due to theincreasing velocity of the piston and increased flow through valve DV,but the pressure does not ordinarily approach the initial peak pressure.When port TP reaches port SP the pressure drops suddenly and deliveryvalve DV closes. Inasmuch as the peak pressure occurs at the beginningof the stroke, irrespective of the adjustment of collar CO, it will beseen that variation of the delivery setting of collar CO by the engineprimary or torque control has no effect on the peak pressure developedwithin the cylinder.

In accordance with the invention, secondary control check valve SV isprovided in FIG. 2b, again with a smaller time constant than deliveryvalve DV, so that valve SV opens briefly during the pressure peak tospill back fuel to the supply, and it will be apparent that increasingengine speed causes greater impulses to valve SV. thereby spilling backmore fuel. Because the magnitude of the pressure peaks does not tend tovary with the delivery setting, it is in general less necessary to use acam to vary the spring loading on valve SV in FIG. 2b. However, the useof such a cam, in the same manner as in FIG. 2a, allows one to moreeasily tailor the secondary control valve spill-back amount to a givenvolu metric efficiency versus speed characteristic, and the use of sucha cam with the pump of FIG. 2b is within the scope of the invention.

While FIG. 2b illustrates a reciprocating variabledelivery pump using aconstant stroke length, its peak pressure characteristic is essentiallythe same as that of a number of reciprocating varible-delivery pumps inwhich the amount of fuel pumped per stroke is varied by varying the pumpstroke length. In such pumps, the peak pressure ordinarily occurs at ornear the beginning of the stroke, and the magnitude of the peak pressuredoes not vary appreciably with the fuel delivery or stroke lengthadjustment. It will be apparent that a secondary control check valve maybe connected to the chamber of such a pump in the same manner as withthe pump of FIG. 2b, with the check valve spring loading being eithervaried or not varied as a function of the primary control or strokelength setting.

FIG. 2c diagrammatically illustrates a third form of reciprocatingvariable delivery, constant stroke-length injection pump ofa type shownin greater detail in FIG. I and also described in detail in my priorapplication Ser. No. 786,233. Piston P is reciprocated by the enginewith some approximation of simple harmonic motion. A passageway withinpiston P communicates with the pump chamber and selectively communicateswith inlet port IP and outlet port OP. The passageway edge positionsvary angularly about the piston so that rotation of the piston variesthe time during a given stroke at which inlet port IP is closed off andthe time at which outlet port OP is opened. thereby varying the amountof fluid pumped during a rightward pumping stroke. The engine primarycontrol rotatably adjusts piston P to vary pump delivery rate. At theleftward position of the piston inlet port IP is fully opened, and atthe rightward end of the pumping stroke outlet port OP is fully opened.The passageway geometry is arranged relative to the two ports so thatoutlet port OP always opens slightly before inlet port IP is completelyclosed off at any angular position of the piston. With inlet port IPclosing outlet port OP is opening, the maximum restriction to flow fromthe pump cylinder occurs during the overlap condition when both portsare slightly open.

Consider initially the operation of the pump of FIG. 20 withoutsecondary control check valve SV. As the piston begins a rightwardpumping stroke. fluid is expelled through fully open inlet port IP, andthe pressure within the pump cylinder remains low. As the inlet portbegins to close off and the maximum flow restriction condition isapproached, the pressure in the pump cylinder increases very rapidly,and then as the maximum restriction overlap condition is passed andoutlet port OP is opened wider, the pressure decreases. The magnitude ofthe peak pressure developed in the cylinder will be seen to depend uponboth engine speed, which determines the flow rate out of the pumpcylinder, and upon the minimum total open area of the two ports whenboth are slightly open. As the piston is rotated to vary the delivery,it will be seen that the time during the stroke at which the maximumrestriction condition occurs will vary. and if the same maximumrestriction condition occurs at different piston velocities. whichprovide different flow rates from the cylinder. it may be seen that thepeak pressure obtained will also vary with the engine primary controlsetting. If the same maximum restriction condition, i.e., same minimumopen area during overlap, is made to occur for all delivery settings,the peak pressure at a given pump drive speed will be seen to beobtained if the maximum restriction condition occurs substantially atmidstroke, when piston linear velocity is greatest, so that the peakpressure for a given drive speed will occur when the pump is adjusted topump approximately one-half of its maximum delivery per stroke.

In accordance with the invention, secondary control check valve SV isconnected from the pump cylinder to spill fuel during the occurrence ofthe pressure peaks. One advantage of the pump of FIG. 2c over those ofFIGS. 2a and 2b is that the delivery check valve DV may be very lightlyloaded, since delivery cannot begin until output port OP is opened,irrespective of pump speed and delivery setting. Furthermore,

while the peak pressure impulse developed in the pumps of FIGS. 2a and2b and applied to their secondary control check valves is limited by theopening of their delivery check valves, the peak pressure developed inthe pump of FIG. 2c is substantially independent of the delivery checkvalve loading, and thus the secondary control check valve used in thearrangement of FIG. need not have a shorter time constant than that ofthe delivery check valve or otherwise be adjusted relative to any othercheck valve.

The above description of FIG. 2c assumes that the same maximumrestriction condition occurs during the overlap condition at all angularadjustments of the piston. By suitably shaping and/or slanting the edgesof the ports relative to the piston passageway edges one can cause thearea of the maximum restriction to vary as the piston is rotated toprovide different delivery rates, and hence one can make the amplitudeof pressure peaks occurring in the pump of FIG. 2c either more afunction of, or less a function of, the delivery setting in whatevermanner one chooses. If the minimum port area occurring during overlap iscaused to increase somewhat with delivery setting up to onehalf of pumpcapacity, thereby decreasing the maximum restriction with an increaseddelivery setting up to onehalf pump capacity, (and thereafter todecrease with increased delivery if more than one-half pump capacity isused) the magnitude of the pressure impulses will tend to vary less withdelivery setting. If the minimum port area during overlap is caused tovary roughly as a sine-squared function with the delivery setting, itwill be seen that the magnitude of the pressure peaks occurring at agiven engine speed can be made theoretically independent of the deliverysetting, so that no variation in check valve spring loading withdelivery setting is necessary. Because the pump of FIG. 2a requires thatthe inlet port be fully closed, providing an infinite restriction at alldelivery settings, the magnitude of the pressure impulses occurring insuch a pump varies markedly in accordance with the delivery setting,since the delivery setting determines the time during the stroke, andhence the piston velocity at the time the restriction is imposed. Thepump of FIG. 2c (and FIG. 1), by not providing an infinite restriction,but instead a controllable partial restriction, the minimum area ofwhich can be made to vary with delivery setting, therefore has themarked advantage that the magnitude of the pressure impulses occurringat a given engine speed may be arranged to vary with delivery setting inaccordance with any desired function, or if desired, arranged not tovary appreciably at all.

It has been shown that while the peak pressure developed during apumping stroke at a given engine speed varies with primary controlsetting with the pump of FIG. 2a, with maximum pressure being developedwhen this pump is pumping at roughly one-half its maximum capacity, thatthe peak pressure developed in the pump of FIG. 2b tends to be largelyindependent of the primary control setting, and that the peak pressuredeveloped at a given engine speed in the pump of FIG. 2c may or may notvary appreciably with delivery setting, depending upon whether its portgeometry is arranged to provide a minimum restriction area which varieswith delivery setting. The amount of fuel spilled back by the secondarycontrol check valve of any of the three systems of FIGS. 2a, 2b and 2cvaries with the peak pressure impulse applied to the check valve in amanner dependent upon the check valve passage geometry, as well as uponits inertia and spring loading.

FIG. 1 illustrates in a cross-section view a form of injector pumpdisclosed and described in detail in my abovementioned Application Ser.No. 786,233. with certain modifications made thereto in accordance withthe present invention. The pump is of the basic type described above inconnection with FIG. 2c, but shown adapted for two-cycle engine use topump oil as well as fuel. The pump comprises a generally-cylindricalcen' tral casting having a rear head 121 and a front head 122 boltedthereto by means of bolts (not shown). with a suitable gasket (notshown) preferably provided between each head and the central casting.Shaft 123 rotated by the engine crankshaft carries eccentric cam 127.Rotation of cam I27 reciprocates tappet 81, which is carried in bushing82 with an O-ring seal 83a. The right end of tappet 81 bears against theleft end of piston 130. which reciprocates within sleeve 129a. A spring133, only a portion of which is shown, is inserted between head 122 anda right-end face of piston 130 and operates to return piston 130. Alower gear sector 83 pinned to piston 130 is engaged by upper gearsector 84 pinned to control shaft 131, so that rotation of shaft 131angularly positions piston 131]. Shaft 131 is angularly positioned byaccelerator pedal or throttle control 103 via arm 104 and a suitablemechanical linkage shown merely as a dashed line. Upper gear sector 84is axially wider than lower gear sector 83 so that the gear sectorsremain enmeshed as sector 83 reciprocates with piston 130.

Oil is supplied from an oil supply tank (not shown) to chamber 128 via acheck valve (not shown) and a pipe connection made at 128a on the sideof central casting 120. Fuel is supplied to chamber 164 from fuel tank146 via conduit c. Oil and fuel inlet ports are provided in sleeve 129aat 134 and 136, respectively. and oil and fuel outlet ports are providedat 135 and 137. Oil piston 161 is urged rightwardly against front head122 by inner coil spring 162. Holes drilled in main casting 120 at 142aand 147a connect the outlet ports with longitudinally-extending passagesin which check valves 150 and 151 are located, and plugs 142c, 147cclose the ends of passages 142 and 147a. Check valves 150 and 151 at theoutlet side of the injector pump each communicate with mixing chamber143 provided in front head 122. Two V-shaped grooves are milled acrossthe outer periphery of piston 30 as shown by dashed lines at 138 and139. The bottom of V-groove 138 communicates with oil chamber 155 insidepiston 130, and the bottom of V-groove 139 communicates with fuelchamber situated to the right of piston 130 and partially within piston130. At various axial positions of piston 130 V-groove 138 connectschamber 155 to only chamber 128 via oil inlet port 134, or to both inletchamber 128 via inlet port 134 and to mixing chamber 143 via outlet port135 and check valve 150, or to only mixing chamber 143 via outlet port135 and check valve 150. At corresponding axial positions of piston 130,V-groove 139 connects fuel chamber 160 to only chamber 164 via inletport 136, or to both chamber 164 via inlet port 136 and mixing chamber143 via outlet port 137 and check valve 151, or to only mixing chamber143 via outlet port 137 and check valve 151. Inlet ports 134 and 136 andoutlet ports 13S and 137 each comprise an opening; which extendspartially around sleeve 129a, with each such slot having a uniformdimension measured in the axial direction of sleeve 129a.

The cutting of V-shaped grooves on the periphery of cylindrical piston130 gives the grooves a width which varies with the angular position ofthe groove around the piston. As is described in greater detail in mymentioned prior application Ser. No. 786.233, varying the angularposition of the piston within sleeve 129a by means of control shaft 131varies the time during a given piston stroke at which the V-grooves willcommunicate with the outlet ports and the time at which the V-grooveswill be cutoff from the inlet ports, and hence determines the amount offuel and oil which the pump will pass to the mixing chamber during thepiston stroke.

Piston 130 is shown at its leftmost position in FIG. 1. As piston 130 isurged rightwardly on a pumping stroke, at the beginning of the strokeV-groove 138 connects oil piston chamber 155 via inlet port 134 tochamber 128 so that oil within chamber 155 is expelled from chamber 155back into chamber 128, and V- groove 139 connects fuel chamber 160 viainlet port 136 to fuel chamber 164, so that fuel is expelled fromchamber 160 back into chamber 164. At an intermedi ate time during thestroke determined by the angular position-of piston 130, the V-groovesfirst reach and unblock outlet ports 135 and 137 and then move out ofcommunication with and block inlet ports 134 and 136. Provision of suchan overlap condition with the outlet ports always slightly openingbefore the inlet ports are fully closed prevents damage due to fluidblockage. Thereafter during the rightward pumping stroke, as the inletports fully close and the outlet ports increasingly open, oil isexpelled from chamber 155 via outlet port 135, and fuel is expelled fromfuel chamber 160 via outlet port 137, and the fuel and oil mix in mixingchamber 143. The mixing chamber connects to a nozzle (not shown) whichinjects the fuel-oil mixture into the engine air intake duct. Asmentioned in my prior application, the fuel and oil are not mixed in amixing chamber in some applications, and instead, only the fuel is pipedto the nozzle and the oil is pumped to various oil holes at desiredlubrication points within the engine.

The basic pump of FIG. 1 and FIG. 2c differs markedly from many somewhatsimilar prior art fuel metering pumps in that an inlet port is closedand a separate outlet port is opened during a pumping stroke, while theprior art generally has left each pump chamber in constantcommunication. with an outlet check valve during the entire pumpingstroke, so that forward pumping past a prior art check valve occurseither immediately (as in FIG. 2b) or as soon as an inlet port is closedoff to prevent return pumping (as in FIG. 2a). Ifthe fluid supply haspositive pressure, the check valve in such prior systems must be loadedto at least the same pressure in order to prevent forward pumping priorto complete closure of the inlet port. And even if the fluid supply isnot pressurized, the pressure in the prior art pump chambers necessarilybuilds up prior to complete closure of their inlet ports, in amountsdependent upon pump speed and dependent upon the amount of restrictionto return flow between the pump chamber and the fluid supply, with theamount of said restriction increasing from a basic amount to completeblockage as the inlet port is gradually closed off. If forward pumpingis not to occur prior to complete closure of the inlet port. the checkvalve in the prior systems must be loaded to the highest such pressurewhich may occur prior to inlet port closure. The heavier check valveloading necessarily results in higher pressures in the pump chamber.thereby requiring a more precise piston-cylinder fit. In the pump ofFIGS. 1 and 2c. forward pumping cannot occur prior to opening of anoutlet port, irrespective of whether the supply is pressurized, andhence the instant at which forward pumping begins during a pumpingstroke remains substantially independent of pump speed and outlet checkvalve loading, making the quantity of fluid delivered per strokesimilarly independent of pump speed and check valve setting.

In accordance with the embodiment of the present invention illustratedin FIG. 1, fuel chamber 160 is connected to fuel inlet chamber 164 via aspring-loaded check valve 163, the spring loading of which is shown madevariable as a function of control shaft 131 position, by means of cam131a carried on control shaft 131. Rotation of control shaft 131, as bymeans of accelerator pedal 103 and arm 104, so as to rotate piston toincrease oil and fuel flow rates causes cam 131a to vary the springloading on check valve 163. The precise shape of cam 131a will dependupon the desired variation of fuel-air ratio, the variation in air flowwith engine speed due to the engine air intake structure, the variationof pump cylinder peak pressure with engine speed, the variation of pumpcylinder pressure with delivery setting, and the variation in the amountof fuel spilled back through check valve 163 with peak pressure, all ofwhich determine the variation in the amount of fuel spilled back for agiven engine speed with a given primary control delivery setting. Insome embodiments of the invention, the spring loading of check valve 163need not be varied as a function of throttle position. In thoseembodiments cam 131a may be eliminated and check valve 163 held inposition with a fixed spring loading by a plug in head 122. Thepassageway which includes a check valve 163 extends generally in adirection so as to intersect shaft 131 if cam 131a is used. If no cam isused it will be apparent that the passageway may extend out radially inanother direction, such as perpendicularly to the plane of FIG. 1.

In the pump of FIG. 1 the inlet and outlet ports are spaced relative totheir respective V-grooves so that maximum restriction to flow from eachV-groove occurs during the intermediate or overlap interval when eachV-groove slightly communicates with both its inlet port and its outletport. Therefore, the maximum pressure which occurs in pump cylinderduring a pumping stroke occurs during that intermediate or overlapinterval when both inlet port 136 and outlet port 137 are both onlyslightly open, so that maximum restriction to flow from chamber 160 isprovided, and the magnitude of the maximum pressure will be seen to varywith pump piston speed. The pressure in chamber 160 will be seento dropfrom its maximum value as piston 130 thereafter continues to travelrightwardly and outlet port.l37 is increasingly unblocked.

As was explained above in connection with FIG. 2c, the maximum peakpressure developed in the pump cylinder for any given engine speed tendsto occur if the maximum restriction condition when both inlet and outletports are slightly open occurs when the piston has maximum linearvelocity. Maximum piston velocity usually occurs somewhere nearmidstroke if an approximation of simple harmonic motion is used toreciprocate the piston. and adjustment of the pump to cause the twoports to overlap around the midstroke causes the pump to operate atapproximately one-half its maximum capacity. If one-half or less of thepump maximum capacity is sufficient to supply the maximum fuelrequirements of the engine, the overlap will occur during the last halfof the pumping stroke, and if the shape and spacing of ports 136 and 137and V-groove 139 provide the same minimum area restriction piston 130 isrotated to give different delivery rates, increasing the primary controlsetting to call for increased delivery will increase the peak pressuredeveloped for a given engine speed and tend to increase the amount offuel spilled back by the secondary control check valve 163, and in suchan arrangement cam 163 may be shaped to provide an increase in checkvalve spring loading as the primary control setting is adjusted toprovide greater fuel flow. If, on the other hand, the maximum fuelrequirements of the engine require more than one-half pump capacity, sothat forward pumping is sometimes required during the first half of thestroke, and the port and V-groove geometry again provides the sameminimum area restriction at different angular po sitions of piston 130,the cam may be shaped to increase check valve spring loading until theprimary control is adjusted to the midstroke overlap condition, andthereafter to decrease the spring loading as greater amounts of fuel arecalled for. However, if rotation of piston 130 is arranged to vary theminimum area of the maximum restriction which occurs during the overlapcondition, the maximum pressure developed in cylinder 160 can be made tovary directly with delivery setting, or not to vary appreciably withdelivery setting, or even to very inversely with delivery setting, ifdesired. If the maximum pressure does not vary appreciably with deliverysetting, it will be apparent that variation of the spring loading oncheck valve 163 becomes unnecessary.

FIG. 1b contains three unrolled or developed views illustrating thegeometry of V-groove 139 relative to ports 136 and 137. Angularadjustment of piston 130 to provide different delivery rates amounts tovertical displacement of the V-groove in FIG. 1b relative to the ports.On each pumping stroke Vgroove 139 moves rightwardly relative to theports from a beginning position in which the V-groove is centered on theinlet port 136. V-groove 139 is shown at I in a minimum deliveryposition at the time during the overlap condition when it leastregisters with inlet port 136 and outlet port 137, at 11 in a mediumdelivery position at the time during the overlap condition when it leastregisters with the ports, and at III in a maximum delivery position atthe time when it least registers with the ports. It will be seen thatthe minimum overlap area varies froma small area in I, to a larger areain II, and then to a smaller area at Ill. Times t,, t; and t indicatethe times after the beginning of the pumping stroke at which the maximumrestriction occurs under the three different delivery conditions. Thusit will be seen that the amount of maximum restriction varies from aminimum at low delivery rates up to a maximum at approximately one-halfca pacity, down to a minimum at maximum delivery. Since piston speed atthe time of the overlap condition varies in approximately the samemanner, it will be apparent that the variation in restriction may beused to offset the variation in piston speed at the time of overlap, sothat the magnitude of the pressure impulses developed at a given enginespeed and pump drive speed tends to be largely independent of the pumpdelivery setting.

If cam 131a is eliminated and a constant spring load is used on checkvalve 163. and if the port geometry provides the same minimumrestriction at different delivery settings, the amount of fuel spilledback through the check valve at a given engine and pump speed willincrease with the primary control delivery setting as the primarycontrol is varied from minimum flow to onehalf pump capacity, therebyleaning out the fuelair ratio. and as the primary control deliverysetting is further advanced at the same engine speed to provide greaterflow than one-half pump capacity, the amount of fuel spilled backthrough the check valve will decrease, thereby providing anincreasingly-enriched mixture at increasing delivery settings.

In the arrangement shown in FIG. 1, wherein increasing fuel spillbackoccurs at increasing engine speeds due to light load conditions but nocomparable oil spillback occurs it will be seen that the amount of oilpumped per stroke remains substantially constant, thereby providinglarger oil-to-fue] and oil-to-air ratios during higher speed-lighterload conditions. Such operation is wholly satisfactory for manytwo-cycle engine applications, and particularly in those two-cycleengine applications where the oil is not mixed with the fuel but insteadpumped to various lubrication joints within the engine.

If desired, the oil-to-fuel and oil-to air ratios may be tailored byproviding a secondary control oil check valve in similar fashion tospill back oil in amounts varying with engine speed.

While FIG. 1 illustrates a system which dispenses metered amounts of oilas well as fuel, such as is used with two-cycle engine systems, it isimportant to recognize that the invention is in no way restricted tofuel injec tion systems which dispense two fluids, and is quite asapplicable to four-cycle engine systems wherein oil is not injected intothe engine.

While the mixing and metering pump of FIG. 1 uses separate oil and fuelpistons (161 and to pump oil and fuel with a desired ratio, analternative embodiment shown in FIG. 1a dispenses with the need for aseparate oil piston, and the need for V-groove 138 on piston 130 and theneed for oil inlet and outlet ports 134 and 135 in sleeve 129a. In FIG.1a oil is supplied to oil chamber 128 via an inlet conduit 128a whichcarries duckbill check valve 601. As cam 127 moves fuel piston 130 on arightward fuel-pumping stroke, thereby increasing the volume of chamber128, oil is drawn into chamber 128 through check valve 601. As spring133 moves piston 130 leftwardly, thereby decreasing the volume ofchamber 128, oil is expelled past oil outlet check valve to mixingchamber 143. The amount of oil which is drawn into chamber 128 during arightward stroke and dispensed to the mixing chamber on the pistonreturn stroke depends upon the cross-sectional area of piston 130 timesthe length of the piston stroke, less the cross-sectional area of tappet81 times the same stroke length, since tappet 81 increasingly enterschamber 128 from bushing 82 as piston 130 increasingly leaves chamber128. If tappet 81 is very slightly less in diameter than piston 130,very little oil will be pumped compared to the amount of fuel pumped. Itwill be seen that a constant amount of oil will be pumped per stroke,irrespective of the adjustment of control shaft 131. In a variety ofsystems. and in particular those which drive constant loads. it isconsidered unnecessary to maintain a constant fuel-oil mixture ratio.The pump in FIG. la is shown without the cam 131a and check valve 163utilized in FIG. 1 to provide secondary control, and such a featureobviously can be added to FIG. la, if desired.

lt will become immediately apparent to those skilled in the art that themain principles of the invention are readily applicable to systems usingvariable stroke length pumps, for example, as well as the specific pumptypes shown. A variety of modifications will occur to those skilled inthe art. Various of the check valves shown as needle valves may compriseball or poppettype valves instead, for example.

It will thus be seen that the objects set forth above, among those madeapparent from the preceding description, are efficiently attained, andsince certain changes may be made in the above constructions withoutdeparting from the scope of the invention, it is intended that allmatter contained in the above description or shown in the accompanyingdrawings shall be interpreted as illustrative and not in a limitingsense.

The embodiments of the invention in which an exclusive property orprivilege is claimed are defined as follows:

l. A fluid pump, comprising, in combination: means defining a pumpchamber having first and second ports; means for supplying fluid to saidchamber; means for cyclically opening and closing said ports to providethree distinct conditions including a first condition in which saidchamber communicates solely with said first port, a second condition inwhich said chamber communicates decreasingly with said first port andincreasingly with said second port, and a third condition in which saidchamber communicates solely with said second port, said ports beingarranged so that the port area communicating with said chamber reaches aminimum finite amount during said second condition and the pressure insaid chamber reaches a maximum during said second condition; and checkvalve means connected to release fluid from said chamber when saidpressure exceeds a predetermined value, said means defining said pumpchamber comprising a cylinder having first and second ports and saidmeans for cyclically opening and closing said ports comprises a pistonadapted to reciprocate within said cylinder.

2. A pump according to claim 1 in which said piston is adapted toreciprocate through fixed-length forward and return strokes within saidcylinder and includes an edge which varies in axial position angularlyaround a portion of said piston so that rotation of said piston withinsaid cylinder varies the times during a forward stroke at which saidports are opened and closed.

3. A pump according to claim 1 in which said piston is adapted toreciprocate through fixed-length forward and return strokes within saidcylinder and is rotatably adjustable to vary the amount of fluid pumpedper stroke through one of said ports by varying the point during astroke at which said one of said ports is opened, and in which saidpiston is shaped so that rotation of said piston varies the minimum portarea reached during said condition.

4. A reciprocating pump, comprising, in combination: a cylinder; pistonmeans adapted to reciprocate within said cylinder; drive means forreciprocating said piston at a velocity which varies over the stroke ofthe piston at a given speed of said drive means. said cylin-' der havingan inlet port and an outlet port. said piston means being operable toincreasingly open said outlet port and increasingly close said inletport during a portion of said stroke and thereby pass through a maximumrestriction condition in which a minimum total area of both ports isopen, said piston means being adjustable to control the amount of fluidexpelled from said outlet port during said stroke by controlling thepoints during said stroke at which said piston means opens said outletport and closes said inlet port. said piston means and said ports beingshaped and spaced so that adjustment of said piston to control theamount of fluid expelled varies the minimum open port area occurringduring said maximum restriction condition. thereby varying the maximumpressure which occurs in said cylinder during said stroke.

5. A pump according to claim 4 having pressuresensitive valve meansoperable when the pressure in said cylinder exceeds a predeterminedvalue to divert fluid from said cylinder and thereby further control theamount of fluid expelled from said outlet port during said stroke.

6. A pump according to claim 4 in which said piston includes an edgewhich varies in axial position angularly around a portion of said pistonso that rotation of said piston within said cylinder controls saidpoints during a stroke at which said piston opens said outlet port andcloses said inlet port.

7. A pump according to claim 4 wherein said drive means reciprocatessaid piston with an approximation of simple harmonic motion.

8. A pump according to claim 4 in which adjustment of said piston toincrease the amount of fluid expelled from said outlet port increasessaid minimum open port area occurring during said maximum restrictioncondition.

9. A pump according to claim 4 in which adjustment of said piston toincrease the amount of fluid expelled from said outlet port decreasessaid minimum open port area occurring during said maximum restrictioncondition.

10. A pump according to claim 4 wherein said drive means reciprocatessaid piston at a velocity which varies substantially sinusoidally overthe stroke of said piston.

1 1. A pump according to claim 4 in which adjustment of said pistonmeans to vary the amount of fluid expelled from said outlet port variessaid minimum open port area occurring during said maximum restrictioncondition in relation to the velocity of said piston when said maximumrestriction condition occurs. so that said maximum pressure issubstantially independent of said adjustment of said piston means.

12. A pump according to claim 4 in which adjustment of said piston meansto vary the amount of fluid expelled from said outlet port varies saidminimum open port area occurring during said maximum restrictioncondition in accordance with a function of the position during a strokeat which said maximum restriction condition occurs, so that the saidmaximum pressure occurring in said cylinder during said maximumrestriction condition varies substantially as a function of said givenspeed of said drive means and does not vary substantially with saidadjustment of said piston means.

13. A pump according to claim 4 in which adjustment of said piston tovar the amount of fluid expelled from Said Outlet P ar s minimalism!cnsttaost areaoc curring during said maximum restriction condition, froma finite first amount of minimum total open port area at a firstdelivery condition in which a first amount of fluid is expelled fromsaid outlet port during each said stroke to a second amount of minimumtotal open port area at a second delivery condition in which a secondamount of fluid is expelled from said outlet port. during each saidstroke.

14. A pump according to claim 13 in which said second amount of fluidexceeds said first amount of fluid and said second amount of minimumtotal open port area exceeds said first amount of minimum total openport area.

15. A pump according to claim 4 in which adjustment of said piston tovary the amount of fluid expelled from said outlet port varies saidminimum total open port area occurring during said maximum restrictioncondition from said second amount of minimum total open port area atsaid second delivery condition to a third amount of minimum total openport area at a third delivery condition in which a third amount of fluidis expelled from said outlet port during each said stroke.

16. A pump according to claim 13 in which said second amount of fluidexceeds said first amount of fluid and said first amount of minimumtotal open port area exceeds said second amount of minimum total openport area.

17. A pump according to claim 15 in which said second amount of fluidexceeds said first amount of fluid and said third amount of fluidexceeds said second amount of fluid and said second amount of minimumtotal open port area exceeds both said first amount of minimum totalopen port area and. said third amount of minimum total open port area.

18. A pump according to claim 15 in which said second amount of fluidexceeds said first amount of fluid and said third amount of fluidexceeds said second amount of fluid, and said second amount of minimumtotal open port area is less than said first amount of minimum totalopen port area and said third amount of minimum total open port area.

1. A fluid pump, comprising, in combination: means defining a pumpchamber having first and second ports; means for supplying fluid to saidchamber; means for cyclically opening and closing said ports to providethree distinct conditions including a first condition in which saidchamber communicates solely with said first port, a second condition inwhich said chamber communicates decreasingly with said first port andincreasingly with said second port, and a third condition in which saidchamber communicates solely with said second port, said ports beingarranged so that the port area communicating with said chamber reaches aminimum finite amount during said second condition and the pressure insaid chamber reaches a maximum during said second condition; and checkvalve means connected to release fluid from said chamber when saidpressure exceeds a predetermined value, said means defining said pumpchamber comprising a cylinder having first and second ports and saidmeans for cyclically opening and closing said ports comprises a pistonadapted to reciprocate within said cylinder.
 2. A pump according toclaim 1 in which said piston is adapted to reciprocate throughfixed-length forward and return strokes within said cylinder andincludes an edge which varies in axial position angularly around aportion of said piston so that rotation of said piston within saidcylinder varies the times during a forward stroke at which said portsare opened and closed.
 3. A pump according to claim 1 in which saidpiston is adapted to reciprocate through fixed-length forward and returnstrokes within said cylinder and is rotatably adjustable to vary theamount of fluid pumped per stroke through one of said ports by varyingthe point during a stroke at which said one of said ports is opened, andin which said piston is shaped so that rotation of said piston variesthe minimum port area reached during said condition.
 4. A reciprocatingpump, comprising, in combination: a cylinder; piston means adapted toreciprocate within said cylinder; drive means for reciprocating saidpiston at a velocity which varies over the stroke of the piston at agiven speed of said drive means, said cylinder having an inlet port andan outlet port, said piston means being operable to increasingly opensaid outlet port and increasingly close said inlet port during a portionof said stroke and thereby pass through a maximum restriction conditionin which a minimum total area of both ports is open, said piston meansbeing adjustable to control the amount of fluid expelled from saidoutlet port during said stroke by controlling the points during saidstroke at which said piston means opens said outlet port and closes saidinlet port, said piston means and said ports being shaped and spaced sothat adjustment of said piston to control the amount of fluid expelledvaries the minimum open port area occurring during said maximumrestriction condition, thereby varying the maximum pressure which occursin said cylinder during said stroke.
 5. A pump according to claim 4having pressure-sensitive valve means operable when the pressure in saidcylinder exceeds a predetermined value to divert fluid from saidcylinder and thereby further control the amount of fluid expelled fromsaid outlet port during said stroke.
 6. A pump according to claim 4 inwhich said piston includes an edge which varies in axial positionangularly around a portion of said piston so that rotation of saidpiston within said cylinder controls said points during a stroke atwhich said piston opens said outlet port and closes said inlet port. 7.A pump according to claim 4 wherein said drive means reciprocates saidpiston with an approximation of simple harmonic motion.
 8. A pumpaccording to claim 4 in which adjustment of said piston to increase theamount of fluid expelled from said outlet port increases said minimumopen port area occurring during said maximum restriction condition.
 9. Apump according to claim 4 in which adjustment of said piston to increasethe amount of fluid expelled from said outlet port decreases saidminimum open port area occurring during said maximum restrictioncondition.
 10. A pump according to claim 4 wherein said drive meansreciprocates said piston at a velocity which varies substantiallysinusoidally over the stroke of said piston.
 11. A pump according toclaim 4 in which adjustment of said piston means to vary the amount offluid expelled from said outlet port varies said minimum open port areaoccurring during said maximum restriction condition in relation to thevelocity of said piston when said maximum restriction condition occurs,so that said maximum pressure is substantially independent of saidadjustment of said piston means.
 12. A pump according to claim 4 inwhich adjustment of said piston means to vary the amount of fluidexpelled from said outlet port varies said minimum open port areaoccurring during said maximum restriction condition in accordance with afunction of the position during a stroke at which said maximumrestriction condition occurs, so that the said maximum pressureoccurring in said cylinder during said maximum restriction conditionvaries substantially as a function of said given speed of said drivemeans and does not vary substantially with said adjustment of saidpiston means.
 13. A pump according to claim 4 in which adjustment ofsaid piston to vary the amount of fluid expelled from said outlet portvaries said minimum total open port area occuring during said maximumrestriction condition, from a finite first amount of minimum total openport area at a first delivery condition in which a first amount of fluidis expelled from said outlet port during each said stroke to a secondamount of minimum total open port area at a second delivery condition inwhich a second amount of fluid is expelled from said outlet port duringeach said stroke.
 14. A pump according to claim 13 in which said secondamount of fluid exceeds said first amount of fluid and said secondamount of minimum total open port area exceeds said first amount ofminimum total open port area.
 15. A pump according to claim 4 in whichadjustment of said piston to vary the amount of fluid expelled from saidoutlet port varies said minimum total open port area occurring duringsaid maximum restriction condition from said second amount of minimumtotal open port area at said secoNd delivery condition to a third amountof minimum total open port area at a third delivery condition in which athird amount of fluid is expelled from said outlet port during each saidstroke.
 16. A pump according to claim 13 in which said second amount offluid exceeds said first amount of fluid and said first amount ofminimum total open port area exceeds said second amount of minimum totalopen port area.
 17. A pump according to claim 15 in which said secondamount of fluid exceeds said first amount of fluid and said third amountof fluid exceeds said second amount of fluid, and said second amount ofminimum total open port area exceeds both said first amount of minimumtotal open port area and said third amount of minimum total open portarea.
 18. A pump according to claim 15 in which said second amount offluid exceeds said first amount of fluid and said third amount of fluidexceeds said second amount of fluid, and said second amount of minimumtotal open port area is less than said first amount of minimum totalopen port area and said third amount of minimum total open port area.